Thermosyphon-powered jet-impingement cooling device

ABSTRACT

A thermosyphon-powered jet-impingement cooling device delivering superior thermal energy dissipation for compact heat sources such as electronic devices. Thermal energy from a heat source travels through the heat source/heat spreader plate interface to the heat spreader plate and heat spreader plate extended surface. Thermal energy is transferred by convection to a single or two-phase coolant media. The heated and/or boiling, less dense coolant begins to expand and rise. The rising coolant or vapor approaches a cold plate and velocity slows due to the greater cross-sectional flow area. The coolant heat energy is released by convection or condensation to the heat dissipation/fluid interface surface, and is then conducted through the cold plate, across the cold plate/heat dissipating device thermal interface, and then to the heat dissipating device. As the cooled media contracts and/or condenses to form droplets, the coolant or droplets begin to fall. As heated or boiling coolant continues to rise, the falling coolant or droplets are both pushed from the high pressure (heated) annulus and pulled from the low pressure (cooled) center through an impingement jet orifice. The coolant impinges against a concave heat/fluid interface surface. The impinging jet of coolant media of the present invention greatly reduces the thermal boundary layer at the point of impingement.

This application is a continuation of application No. 08/758,025, filedNov. 27, 1996, now U.S. Pat. No. 5,864,466.

BACKGROUND--FIELD OF INVENTION

The present invention relates to a device for cooling by naturalconvection, with or without phase change. Specifically to such devicesincorporating unique surface contours that provide a thermosyphon effectwith jet-impingement cooling of the heat source.

BACKGROUND--DESCRIPTION OF PRIOR ART

In many fields, the control and dissipation of excess thermal energy isdesired. The speed and reliability of modern electronic equipment dependon the cost-efficient dissipation and control of waste heat. Ascomponent performance increases and size decreases, the heat flux(power/area) spirals upward.

Air has been used as a coolant since the origin of electronic devices.Gravity-dependant, natural air convection produces a heat transfercoefficient of about 5 to 30 W/m² °C. Using only natural convection withair, a typical 1 Watt, 45 millimeters per side, horizontally-mounteddevice, has a temperature rise (ΔT) of roughly 45° C. Forced airconvection (3 meters per second velocity) can reduce the device's ΔT toroughly 15° C. As electronic components have required better heatdissipation techniques, designers have increased the power of aircooling systems. High velocity air systems reach a point of diminishingreturns. Higher power cooling systems also generate their own high wasteheat loads and high noise levels from air moving devices. These highheat loads and high noise levels must then also be controlled. The costconstraints in this approach are nowhere more apparent than in the fieldof personal computers. These systems use high performance VLSI (VeryLarge Scale Integration). CMOS (Complementary Metal-OxideSemiconductor), and RISC (Reduced Instruction Set Computer)technologies. Some high heat flux electronic devices present a thermalobstacle that forced convection with air cannot overcome. Therefore,there is a great need to reduce an electrical device's temperature inthe most efficient manner possible. CMOS devices especially, show greatincreases in performance and reliability when operated at reducedtemperatures. Tobey teaches the use of reduced temperatures to gainincreased reliability and speed in CMOS devices in U.S. Pat. No.4,812,713 (1989).

Various methods enhance the heat transfer capabilities of forced airconvection systems. Such methods are well known in the art and includeenhanced surfaces, swirl flow, jet-impingement, vibration, injection orsuction, and electrostatic or magnetic fields. These techniques arebeneficial, but are limited by the physical properties of the ambientair, and can require significant mechanical apparatus.

Force air convection heat transfer coefficients range from approximately30 to 300 W/m² °C. An arrangement for air-impingement cooling ofelectronic modules is disclosed by Dunn et al. in U.S. Pat. No.4,277,816 (1981). Dunn utilizes air, directed under pressure to theelectronic modules. This device lacks a leak-proof air exhaust passages,therefore, the cross-flow degradation of the cooler inlet air withheated exhaust air causes a great loss of efficiency. Modules at thesystem air exhaust are exposed to exhaust air that has been heated bythe previous modules. Hwang et al., in U.S. Pat. No. 4,233,644 (1980)teaches the use of a pair of air-moving devices to pull air through aplenum chamber. The air is then directed to impinge against theelectronic modules. In Hwang's device, the air is directed to separateinlet and exhaust passages. In most instance, the use of this devicerequires a major level of redesign of the entire electronic cabinet, orthe intention to use such a cooling scheme while designing theequipment. Hamadah et al., in U.S. Pat. No. 5,063,476, discloses anapparatus for controlled air-impingement cooling. This device utilizescontrolled inlet and exhaust air channels contained in circuitboard-sized modules. Hamadah's apparatus is thought to provide a lowerlevel of redesign for an overheating electronic cabinet, but stillsuffers from the physical properties of air and the limitations ofair-moving devices. U.S. Pat. No. 4,682,651, by Gabuzda (1987),describes a heat sink array for use with air-impingement systems. Thenovelty of this device is that the heat sink is segmented to minimizethe effect of thermal contraction and expansion on the circuit packages.Thermal mismatch is a concern between the circuit board and the leadedside of a circuit package, as this mismatch stresses the solder joints.Gabuzda's device is only beneficial when the circuit package is solarge, and the heat transfer so inefficient, that controlling thethermal mismatch on the non-leaded side of a component will have someeffect on reliability.

Peltier junctions are often suitable for cooling discrete electronicdevices below ambient temperature. The use of Peltier junctions in thismanner is taught in U.S. Pat. Nos. 4,328,759 by Hunsperger, U.S. Pat.No. 4,685,081 by Richman (1987), and U.S. Pat. No. 4,812,733 by Tobey(1989). The obstacle to incorporating a Peltier junction into a designis the resultant condensate and the inefficiency of these devices. Byexample: to cool a 10-Watt device, an additional 25 Watts of heat may bedeveloped by the Peltier junction itself, then requiring removal of 35Watts of heat. Some computer designers have used a Peltier junction tocool the central processing unit (CPU), and then additional air-movingfans to cool the Peltier junction.

In 1935 experimenters reported using de-ionized water to coolcomponents. Natural convection in water can produce heat transfercoefficients in the range from 300 to 2,000 W/m² °C. For the previouslymentioned one Watt electronic device, the ΔT is only 1° C. when immersedin water. Special synthetic dielectric coolants have recently beendeveloped for electronic cooling. Some of these are silicate ester,perfluorocarbon, and polyalphaolephin based compounds.

Many systems, such as the Cray-2 Superconductor, have been designed tooperate immersed in a perfluorocarbon fluid using natural convection.This technique is complicated by the properties of the fluid; highvolatility and low surface tension. These problems have been addressedin the industry by securing the perfluorocarbon fluid in flexibleplastic bags. These bags are attached to the electronic circuit boards.This approach decreases the effectiveness of free convection byrestricting the internal fluid motion, and also adds unwanted thermalimpedance as the heat must pass through the thermally insulative plasticbag containing the fluid.

Forced liquid convection heat transfer coefficients are much higher thanforced air (30 to 300 W/m² °C.), and range from approximately 300 to6,000 W/m² °C. The prior art for liquid jet-impingement devices does nothave the limitations of the properties of air, but still suffer from thedrawbacks associated with liquid moving devices. In my own U.S. patentapplication, 674,820 of 1991, a flexible module is described that allowsthe controlled application of multiple jets of an impinging liquidcoolant. This device is unique because novel features are used to sealthe liquid from the electronic components while still delivering adegree of geometric flexibility, and the maximum amount of heattransfer. Bland et al., in U.S. Pat. No. 4,494,171 (1985), and Crowe, inU.S. Pat. No. 4,901,201 both describe very complex forced impingementcooling apparatus referred to by the acronym CHIC, for Compact HighIntensity Cooler. Crowe's device offers a means to provide dedicatedcooling for high heat flux devices located in the midst of othercomponents requiring minimal cooling. Users of Crowe's invention mustknow the exact location of each of the high heat flux devices in orderto use the CHIC in only that location, because a lower cooling level isprovided to all other components. U.S. Pat. No. 4,750,086, by Mittal(1988), teaches the use of a forced liquid jet-impingement device thatincorporates a bellows seal and a thick heat spreader. In all of theaforementioned jet-impingement devices, liquid and air, a pressurizedmedia is used for forced-convection heat transfer. Forced convectionrequires an external power source to propel the coolant, in addition tothe coolant moving device itself. These devices create waste heat andmust be located to distribute the cooling media in an evenly distributedmanner. This entails the use of plumbed lines and leak-proof modules. Nodevice known, has used the ability of a natural convection thermosyphonto power the jet-impingement apparatus.

Although liquid cooling systems are much more efficient than air, theyhave been cost prohibitive in small systems such as personal computers,because of the necessary fluid-moving device, plumbed lines, andleak-proofing.

Cooling a heat source by means of phase change (boiling and condensing)is the most efficient method used in the industry. While single-phaseforced liquid convection heat transfer coefficients range from 300 to6,000 W/m² °C. boiling water heat transfer coefficients range fromapproximately 3,000 to 60,000 W/m² °C. and condensing steam systemsoperate in the range of 6,000 to 120,000 W/m² °C. Mitsuoka, in U.S. Pat.No. 3,986,550 (1976) describes a cooling device operating on the phasechange principle that maintains separation between the liquid and gasphases of the coolant media. No attempt is made in Mitsuoka's device toincrease the heat transfer coefficient by use of internal extendedsurfaces, jet-impingement techniques, or other geometric forms. In fact,in the described patent, Mitsuoka teaches away from impingement by usinga variety of devices placed in the flowstream to prevent intermixing ofthe fluid and vapor phases of the coolant.

Within the prior art of natural convection, air-cooled heat sinks, it isnot evident that a specific geometric form will channel the naturalconvection flowfield currents to gain greater heat dissipation throughminimized thermal boundary layers, or that this geometric form (withadjustment for coolant properties) can be used in conjunction withnatural convection liquid-immersed heat sinks. Although there are manyexamples of prior art single-phase thermosyphon devices, and prior arttwo-phase thermosyphon-powered cooling devices, it is clearly unobviousto those skilled in the art that there are advantages in heat transfercoefficient and operating efficiency when the jet-impingement techniqueis combined with either or both of these two approaches, ie,single-phase thermosyphon powered jet-impingement, or two-phasethermosyphon-powered jet impingement. Further, in somethermosyphon-powered prior art (Mitsuoka), the return of the condensateto the boiling pool is shown as occurring in the geometric center of theheat source. The prior art does not recognize that specific novelgeometric features that connect fluid transfer channels affect the heattransfer coefficient, and that significant departure from theseunobvious geometric ratios will result in a device that may operate in areverse mode, resulting in a severe degradation in performance. A reviewof the literature yields no discussion of these combinations, or thesupposition that such combinations are advantageous or even possible.

OBJECTS AND ADVANTAGES

The present invention transfers heat by use of a self-contained,jet-impingement, thermosyphon effect and does not require large volumesof moving coolant, plumbed lines, or extensive module redesign whenincorporated. The thermal mass of the liquid-filled current inventionretards the dramatic temperature shock caused by applying power to adevice, and greatly enhances solder joint reliability.

The present invention transfers heat by use of a jet-impingement,thermosyphon effect and does not generate or use additional energy.

The present invention, in a single-phase configuration, can use any gasor liquid cooling media that expands when heated. In either asingle-phase or two-phase configuration, it offers significantly higherheat transfer coefficients than simple device immersion in an identicalcooling media under similar cooling parameters.

The present invention can use a perfluorocarbon fluid. Because theliquid is sealed within the cooling device, leakage is eliminated.Thermal impedance is low because the internal heat transfer surfaces aremetallic or another material offering high thermal conductivity.

The present invention provides an efficient and inexpensive liquidcooling system, primarily for the thermal management of small electroniccomponents, but is applicable to any heat generating device.

Therefore, in accordance with an aspect of the present invention,several primary objects and advantages of the present invention are:

(a) to provide a new and improved method and apparatus of naturalconvection heat transfer that is contained in one small module;

(b) to provide an improved natural convection heat transfer device withcertain novel geometric features that provide surprising and highlyadvantageous results, primarily in the form of greater thermal energytransfer coefficients;

(c) to provide a method and apparatus for natural convection heattransfer with the cooling media applied as an internal impingement flowconfiguration that provides the lowest thermal resistance available in alow-cost, passive configuration;

(d) to provide a method and apparatus for leak-proof natural convectioncooling that controls the large thermal volume expansion rate of thecoolant;

(e) to provide a method and apparatus for natural convection coolingthat provides the thermal benefits of forced convection without the needfor an active coolant media moving device;

(f) to provide an improved apparatus with all of the previouslydescribed benefits for natural convection cooling device applied to atwo-phase coolant device.

(g) to provide a method and apparatus for phase change with jetimpingement of the coolant media cooling that provides greater thermalbenefits than phase change devices that do not incorporate specializedcoolant passages;

(h) to provide an integral method with cooling that minimizes thethermal shock experienced by electronic devices when operated at highclock rates under transient conditions.

These and other objects and advantages of the present invention will nodoubt become obvious to those of ordinary skill in the art after havingread the following detailed description of the preferred embodiment,which is illustrated in the various drawing figures.

DRAWING FIGURES

In the drawings, closely related figures have the same number butdifferent alphabetic suffixes.

FIG. 1 shows the present invention as a basic geometric form in anisometric view;

FIG. 2 depicts the present invention in a cut-away, isometric view;

FIG. 3 illustrates the present invention in a liquid impinging-jetembodiment in an external isometric view;

FIG. 4 depicts the preferred embodiment of the present invention in acut-away, isometric view;

FIG. 5 presents a sectional view of the liquid coolant circulation pathof the present invention;

FIG. 6 shows the pin-fin heat sink and heat spreader plate of thepresent invention;

FIG. 7 portrays the device in a sectional view in the phase changecoolant configuration, with spring clip, in a typical application.

FIG. 8a portrays the air flow and thermal boundary layer in proximity toa horizontally oriented, heated device;

FIG. 8b demonstrates the jet-impingement flow and influence on thermalboundary layer reduction.

DESCRIPTION--FIGS. 1 TO 8

FIG. 1 illustrates an external isometric view of a thermosyphon-poweredjet-impingement cooling device 11 of the present invention in the mostbasic form. The thermosyphon-powered jet-impingement cooling device 11is enveloped by a fluid coolant 30. In this embodiment, the fluidcoolant is ambient air. A geometric body form 13 is mounted above a heatspreader plate 12 and is constructed of a thermally insulative materialsuch as polyurethane. A heat spreader extended surface 28 of heatspreader plate 12 allows greater surface area and heat dissipation thana flat plate. A fluid contraction surface 48 leads fluid coolant 30 intoan impingement jet channel 52.

FIG. 2 shows an isometric cut-away view of the geometric form of thepresent invention. A plurality of pin-fins 64 are mounted to heatspreader plate 12 on a heat/fluid interface surface 18. A point ofimpingement 42 on heat spreader plate 12 is at the lowest point of theconcave surface of heat/fluid interface surface 18. Geometric body form13 is formed by a channel wall body 26 having no large internal materialvoids. The underside of geometric form 13 is an angled fluid expansionsurface 46 that contacts and is supported by the uppermost surface ofpin-fins 64 which are constructed of a highly thermally conductivematerial such as copper or aluminum. The inward surface of channel wallbody 26 forms fluid contraction surface 48 that forms impingement jetchannel 52.

FIG. 3 shows an external isometric view of the preferred embodiment ofthe present invention. A curved outer wall 22 forms the outer envelopeof the active portion of the device. Outer wall 22 is primarilyspherical in shape, non-porous, and a good thermal insulator such as anon-filled polymer. A cold plate 14 allows heat transfer from theinternal fluid to the exterior of the apparatus. A heat dissipatingdevice 16 is shown as having a pin-finned extended surface 62. Heatdissipating device 16 and pin-finned extended surface 62 are constructedof a highly thermally conductive material such as copper or aluminum.

A cut-away view of the internal components of the thermosyphon-poweredjet-impingement feature of the present invention is shown in FIG. 4. Aheat source 10 can be any heat generating device. Examples of heatsources are electronic devices, motors, engines, light sources, nucleardevices, etc. In the preferred embodiment of the present invention, heatsource 10 is characterized as a square microprocessor computer chip.Heat spreader plate 12 is constructed in the shape of a square frommaterial having a high value of thermal conductivity. Heat spreaderplate 12 becomes circular as it extends away from square heat source 10and toward circular heat/fluid interface surface 18. Heat spreader plate12 is in intimate thermal contact with heat source 10 evenly conductingthe distributed heat to heat/fluid interface surface 18. If heatspreader 12 is not made as a square to circular shaped transition form,heat spreader 12 should be circular through-out with a diameter as largeas the diagonal distance across square heat source 10. Heat transfer atheat/fluid interface surface 18 is through convection. The center of thesurface of heat/fluid interface surface 18 is directly below impingementjet channel 52 and contains point of impingement 42.

Referring now to FIG. 6, the central area of heat/fluid interfacesurface 18 of heat spreader plate 12 is concave and free from surfaceprotrusions. The concavity of the surface allows the coolant to comeinto closer proximity with heat source 10 and also increases the areaavailable for heat transfer without significantly increasing fluidenergy requirements.

Referring again to FIG. 4, cold plate 14 absorbs heat from fluid coolant30 at a heat dissipation/fluid interface surface 20. Heat spreader plate12 and cold plate 14 are constructed of oxygen-free copper. The highthermal conductivity of cold plate 14 allows the heat to be evenlydistributed to heat dissipating device 16. Heat dissipating device 16can be attached or pressed against an extended surface or such apparatusthat is to direct the thermal energy to the ultimate heat sink (notshown). Heat spreader plate 12 and cold plate 14 are roughly equivalentin diameter to heat source 10 and heat dissipating device 16. If heatspreader plate 12 or cold plate 14 are smaller than heat source 10 orheat dissipating device 16, heat spreader plate 12 and cold plate 14 aresized to a thickness that will allow full heat spreading, but not sothick as to add unnecessary thermal resistance. The actual thicknessesof heat spreader plate 12 and cold plates 14 are dependant on thethermal conductivity of the chosen material and the ratios of equivalentradius (a/b) where: a=equivalent radius of heat source 10, and b=theequivalent radius of heat spreader plate 12, and the equivalent radiusof a square surface is found by (L×W/π)⁵, where L=length of one side,and W=length of side perpendicular to side L. A heat source/heatspreader plate thermal interface 58, and a cold plate/heat dissipatingdevice thermal interface 60 are optimized for thermal conduction by suchmethods as thermal pads, greases, joint compounds, gold-silicon eutecticbonding, or chromium-copper interface layers. Such methods are wellknown in the art and taught by Lee in U.S. Pat. No. 4,620,215 (1986).

Curved outer wall 22 is hermetically sealed to heat spreader plate 12and to cold plate 14. An inner channel wall 24, like inner surface ofouter wall 22 is smooth and free from surface defects and irregularitiesthat may induce fluid turbulence. Inner channel wall 24 forms the outersurface of channel wall body 26. Channel wall body 26 is toric in shape,and is thermally non-conductive through-out. Fluid expansion surface 46comprises the lower surface of channel wall body 26.

Channel wall body 26 is constructed of a smooth-skinned, foamed plastic,such as polyurethane. If the inside of channel wall body 26 contains oneor more voids, these voids must be very small and result in a Rayleighnumber of less than 1600, or decreased performance will result. Channelwall body 26 is held in place by attachment to the uppermost surfaces ofheat spreader extended surface 28.

The invention is filled by fluid coolant 30 that is used to enhance thenatural convection process and has a numerically high, free convectionparameter. Other fluid coolant factors such as temperature range,toxicity, material incompatibility, electrical characteristics, etc.,should also be considered.

Fluid contraction surface 48 of channel wall body 26 is shaped to allowfluid coolant 30 to contract and fall to point of impingement 42 with aminimum loss of momentum and fluid energy. Fluid contraction surface 48has a non-linear curvature The diameter of impingement jet 52 is sizedto provide a maximum velocity with a minimum of fluid energy loss.

Referring now to FIG. 5, fluid expansion surface 46 increases indistance from heat/fluid interface surface 18 as the radial distanceincreases from point of impingement 42. An optimum angle 54 of theincreasing distance is related to the heat input, volumetric flow rate,and the fluid properties of thermal expansion, and specific heat. Theslope of angle 54 is linear. A highest edge of angle 56, of fluidexpansion surface 46 allows the volume between heat/fluid interfacesurface 18 and fluid expansion surface 46 to equal the thermally-inducedincreases in fluid volume at the volumetric flow rate established byheat source 10.

Referring again to FIG. 6, pin-fins 64 are of such an arrangement so asto minimize the fluid energy loss. A pin-fin inner circle 66 containspin-fins 64 that are staggered to, and are of an equal number butsmaller diameter than a pin-fin outer circle 68. The smaller diameter ofpin-fins 64 in pin-fin inner circle 66 allows equal cross-sectional areaand therefore velocity, for fluid flow through pin-fin outer circle 68.

Referring to the preferred two-phase embodiment shown in FIG. 7. Heatsource 10 is in intimate thermal contact with heat spreader plate 12 anda condenser mass 84 is in intimate contact with heat dissipating device16 through heat source/heat spreader plate thermal interface 58 and coldplate/heat dissipating device thermal interface 60 respectively.Condenser mass 84 is constructed of a highly thermally conductivematerial such as copper or aluminum. The intimate contact described isprovided by a spring clip 82 which has a characteristic highly elasticshape. Heat dissipating device 16 is shown with pin-finned extendedsurface 62. Heat source 10 is depicted as a computer microprocessorintegrated circuit mounted to a printed circuit board 86 by a pluralityof electrical signal, power, and grounding, etc., pins 88. Printedcircuit board 86 is usually made of a fiber reinforced material or arigidized flexible dielectric with copper circuit traces attachedthereto. Pins 88 are usually of a gold or tin-plated material such ascopper. The heat/fluid interface surface of previous embodiments hadbeen replaced with a high thermal conductivity and geometricallysimilar, but rough-surfaced, dendritic heat/fluid interface surface 74.The rough surface of dendritic heat/fluid interface surface 74 allowsnumerous sites for the initiation of nucleate pool boiling of atwo-phase fluid coolant 90 to occur. For electronic devices, two-phasefluid coolant 90 may be a perfluorocarbon compound with a maximumboiling point roughly 25° C. below the maximum junction temperature ofthe semiconductor. The dendritic surface 74 may be formed by selectiveetchants, laser cutting, or sintering, or as part of a powdered metalprocess for manufacturing heat spreader plate 12. The powdered metalprocess must ensure that the finished product is non-permeable to avaporized coolant 76 phase of two-phase fluid coolant 90. A boilingliquid phase 72 of two-phase fluid coolant 90 fills this embodiment ofthe current invention to roughly the geometric halfway point. Vaporizedcoolant phase 76 of two-phase fluid coolant 90 fills the remaininginternal void within outer surface 22. The condensate of two-phase fluidcoolant 90 forms a coolant condensate droplet 78 on a condenser surface80 of condenser mass 84. These coolant condensate droplets 78 aregravity-urged to become a falling coolant droplet 92 at the lowest pointof condenser surface 80. The cooler (relative to boiling liquid phase72) falling coolant droplet 92 falls into impingement-jet supply pool 94formed on the perimeter by fluid contraction surface 48 and on thebottom by point of impingement 42. The geometric similarity of dendriticheat/fluid interface surface 74 and heat/fluid interface surface 18 ofthe previous embodiments includes the adoption of heat spreader extendedsurface 28, comprised in this embodiment of pin-fins 64 configured asshown in FIG. 6. Channel wall body 26 is supported by the uppermostsurfaces of pin-fins 64. Channel wall body 26 should be constructed of amaterial that will compress inwardly without unduly disrupting thesmooth surface contour of inner channel wall 24 while maintaining athermally insulative material property.

Referring now to FIG. 8a, heat source 10 is shown mounted in ahorizontal position. A thermal boundary layer 70 characteristic ofnatural convection on horizontal surfaces is shown. The shape of thermalboundary layer 70 is formed by factors of the Reynolds number and othercoolant properties.

Referring now to FIG. 8b, impingement jet channel 52 of the presentinvention is shown in relation to point of impingement 42 and thegreatly reduced thermal boundary layer 70.

OPERATION--FIGS. 1 TO 8

Referring now to FIG. 1 of the present invention, geometric form 13 inthis embodiment is designed to be used in an ambient air environment.When attached to a typical computer chip, geometric form 13 and heatspreader extended surface 28 of the present invention will reduce thetemperature rise of the computer chip from the previously indicated 45°C. to a temperature rise of roughly 25° C.

FIG. 2 depicts an external view of the present invention when theinvention is filled with a liquid coolant media. Liquid coolants have amuch higher numerical free convection parameter, and can thereforeprovide greater heat absorption and dissipation.

Referring to FIG. 5, thermal energy from heat source 10 travels throughheat source/heat spreader plate thermal interface 58 to heat spreaderplate 12. Thermal energy is then conducted through heat spreader plate12 to heat/fluid interface surface 18 and heat spreader extended surface28. Thermal energy is transferred by convection to fluid coolant 30.Heat spreader extended surface 28 enhances the transfer of thermalenergy because of the increased surface area. A heated fluid coolant 32is heated and begins to expand 34. A heated, less dense coolant beginsto rise 36. Rising fluid coolant 36 approaches cold plate 14 andvelocity shows due to the greater cross-sectional flow area. Fluid heatenergy is released by convection to heat dissipation/fluid interfacesurface 20. Heat energy is then conducted through cold plate 14 thenacross cold plate/heat dissipating device thermal interface 60 and thento heat dissipating device 16. A cooled fluid 38 contracts. As cooledfluid 38 contracts, a fluid coolant begins to fall 40. As heated fluidcontinues to rise 36 falling fluid 40 is both pushed from the highpressure side and pulled from the low pressure side through impingementjet orifice 52. Fluid impinges against concave heat/fluid interfacesurface 18 at point of impingement 42.

While natural convection heat transfer will dominate conduction whenthere is a sufficient distance and ΔT between two surfaces (whenimmersed in a gaseous or liquid coolant media), the operating range ofthe jet-impingement feature of this invention is limited by theoptimization of the geometry for a specific coolant media. The envelopefor efficient operation during the two-phase jet-impingement process issmaller still. In some combinations of power input, geometry and coolantmedia viscosity, the jet-impingement process may operate in reverse,lowering the heat transfer capability of the device dramatically. Suchreverse operation has been seen during computer simulations but did notoccur during actual testing.

Referring now to FIG. 8a, horizontally mounted heat source 10 releasesheat through a thick thermal boundary layer 70. The natural convectionof the ambient media causes thermal boundary layer 70 to become greaterat the geometric surface center of heat source 10. The center of heatsource 10 is generally where an electronic device heat source producesthe greatest heat density.

Referring now to FIG. 8b, impingement jet 52 of the present inventiongreatly reduces thermal boundary layer 70 at point of impingement 42.

THEORY OF OPERATION

Reliability losses have been related to an electronic device's operatingcharacteristic changes. Temperature changes in particular are known toinduce failures within a device. The failures attributed to temperaturecan be quantified as a normalized failure rate defined by an Arrheniusequation, θ_(n) =θT/θT_(r) =Exp [(E.sub.Λ /k)(T_(r) ⁻¹ -T⁻¹)]. Whereθ=Failure rate; θ_(n) =Normalized failure rate; T=Absolute junctiontemperature (K); T_(r) =Reference temperature (K); E.sub.Λ =Activationenergy (eV); and k=Boltzmann's constant: 8.616×10⁻⁵ (eV/K). The heattransfer path through the invention begins at the heat source/heatspreader plate thermal interface 58. The rate of heat transfer iscontrolled by conduction and is influenced by the pressure at theinterface. The contact resistance is affected by many factors, butprimarily by the roughness of the contact surfaces, and the interstitialfluid in the void space. One formula for determining the heat transfercoefficient (h) through heat source/heat spreader plate thermalinterface 58, and cold plate/heat dissipating device thermal interface60 is h=[k_(f) Y/(σ₁ +σ₂)]+8000 k (CP_(n) /H)^(RG). Where k_(f) =thermalconductivity of the interstitial fluid (W/m² K); Y=peak line of contactof surfaces (m); σ₁ =RMS (root-mean-square) roughness of first surface(m); σ₂ =RMS roughness of second surface (m); k=thermal conductivity ofcontact materials (W/m² K); C=constant (dimensionless); P_(n) =apparentpressure (kg/m²); H=hardness (kg/m²).

Optimum angle 54 of the increasing distance from point of impingement 42is related to the heat input Q, volumetric flow rate m, fluid propertiesof thermal expansion β, and specific heat c_(p). Angle 54 is linearbecause of the linear nature of β. Calculating Qβ/mc_(p) results in thevolume increase per unit temperature rise over time. Highest edge 56, offluid expansion surface 46 allows the volume between heat/fluidinterface surface 18 and fluid expansion surface 46 to equal theincrease in fluid volume at the volumetric flow rate established by heatsource 10. If the inside of channel wall body 26 contains one or morevoids, these voids must result in a Rayleigh number (ρ² gL³ βc_(p)ΔT/μk) of less than 1600, where L is the void's greatest dimension and kis the thermal conductivity of the media filling the void, or decreasedperformance will result. Fluid coolant 30's suitability for use as acoolant in the present invention is found by a relative numericallyhigh, free convection parameter, ρ² gβc_(p) /μk. Other fluid coolantfactors such as temperature range, toxicity, material incompatibility,electrical characteristics, etc., must also be considered. Fluidcontraction surface 48 has a non-linear curvature because of the effectof gravity g, on the linear β.

Impingement jet 52 is sized to provide a maximum velocity with a minimumof fluid energy loss. Available pressure generated by the thermosyphoneffect is found by ρβΔTgL. Due to the varying cross-sectional flow areathrough out the device, fluid coolant velocity varies widely. A fluidcoolant reference velocity is found by √βΔTgH, where H is the hydraulicheight of the apparatus.

Impingement jet 52 of the present invention greatly reduces thermalboundary layer 70 at point of impingement 42. The Nusselt number for asubmerged impinging jet process such as the present invention is foundby the empirical formula 1.29 Re⁵ Pr⁴. I believe the convectioncoefficient, h_(c), is based on half the length of heat source 10,because of the radial direction of coolant flow. The Reynolds number ina natural convection device is found by the square root of the Grashofnumber (L³ ρ² gΔT/μ²). In this embodiment d replaces L in the Grashofnumber and is the diameter of impingement jet 52. Actual ΔT is found byQ/(h_(c) Λ_(s)).

Heat dissipation is very efficient in the present invention. In the 1Watt electronic device described earlier, natural convection in ambientair yielded a temperature rise of 45° C. In the present invention withwater as fluid coolant 30, ΔT is calculated as about 2° C. above theimpingement jet, fluid coolant temperature.

For more efficient operation, a coolant media that changes phase withinthe working temperature of this device is advantageous. Generally, byway of example, perfluorocarbon fluids will operate within thetemperature range of electronic devices. Because of the greater internalpressure during phase change, channel wall body 26 will accept theincreased pressure by equally compressing. The compression of the smallvoids within the channel wall body will allow the device to operatewithout exceeding the internal pressure design limits.

When the device is operated in the phase change mode, heat transferthrough boiling and condensation are important. At the initiation ofboiling, the steady state jet impingement operation is disrupted by therandomized motion of the turbulent coolant media. During this phase, thedevice operates in a pool boiling configuration. Heat transfer is foundby the following equation: ##EQU1## where: q"=heat transfer (W/m²),ΔT_(x) =excess temperature (°C.), p=operating pressure (bar), p_(cr)=critical pressure (bar). During operation, the heat transfer from thesurface must be kept below the critical heat flux (CHF). The CHF can befound by: ##EQU2## where q_(CHF) =critical heat flux (W/m²), ρ_(v)=saturated vapor density (kg/m³), ρ₁ =saturated liquid density (kg/m³),h_(fg) =latent heat of vaporization (J/kg), σ=surface tension of theliquid-to-vapor interface (N/m), g=gravitational acceleration (m/s²).

When the saturated vapor comes into contact with the condenser surface,condensation is produced. During normal operation, continuous dropletsof fluid form and fall into the impingement jet channel. Heat transferfrom the condensing vapor is primarily due to conduction through thedroplets. Because the droplets flow, due to gravity, the formula forpredicting the heat transfer is greater than that seen in the standardheat conduction formula, and follows: ##EQU3## where: h_(c) =averageheat transfer coefficient (W/m² °C.), ρ₁ =saturated liquid density(kg/m³), ρ_(v) =saturated vapor density (kg/m³), g=gravitationalacceleration (m/s²), h_(fg) =latent heat of vaporization (J/kg),k=thermal conductivity of condensate (W/m²), ψ=average angle of surfacefrom horizontal ("), μ=absolute viscosity (N s/m², L=vertical surfaceheight, T_(s) =surface temperature (°C.), T_(sv) =vapor saturationtemperature (°C.).

SUMMARY, RAMIFICATIONS, AND SCOPE

Accordingly, the reader will see that the thermosyphon jet-impingementfeature of this invention can be used to cool any device in which theadvantage of enhanced natural convection cooling is important.

Although the description above contains many specifications, theseshould not be construed as limiting the scope of the invention but asmerely providing illustrations of some of the presently preferredembodiments of this invention. For example, the external envelope mayhave other geometric shapes, such as cubic, pyramidal, conic, prismic,cylindrical, etc.; the fluid may be allowed to vaporize upon heating andcondense upon cooling (two-phase); the fluid may be a dielectric andallowed to come into direct contact with an unpackaged electroniccircuit, etc.; the internal and external extended surfaces may haveother shapes such as lanced, offset, straight, wavy, etc.; heat transferto the ultimate heat sink may take place at a remote location, wherebythe coolant is thermosyphoned to the remote location for cooling andback again, etc.

Thus the scope of the invention should be determined by the appendedclaims and their legal equivalents, rather than by the examples given.

I claim:
 1. A natural convection heat transferring apparatus fortransferring heat energy from a higher temperature heat source to alower temperature heat sink, said apparatus comprising:a heat spreaderplate; a heat sink; and a toric body connected to and positioned aboveand in thermal contact with said heat spreader plate and positionedbelow and in thermal contact with said heat sink to allow a coolantfluid to be heated by said heat spreader plate and flow upwardly alongan outer surface of said toric body and then to be cooled by said heatsink and flow downwardly through a center of said toric body directingthe coolant fluid to impinge on said heat spreader plate and continue tocirculate around and through said toric body due to a gravity dependentthermosyphon effect created by said toric body.
 2. The apparatusaccording to claim 1 wherein said heat spreader plate is connected tosaid toric body by a thermally conductive connection.
 3. The apparatusaccording to claim 1 wherein said heat spreader plate is provided withan upper surface beveled upwardly from a center to an outer edgethereof.
 4. The apparatus according to claim 1 wherein a lower surfaceof said toric body tapers upwardly from a center to an outer edgethereof.
 5. The apparatus according to claim 1 wherein said thermallyconductive connection is defined by at least one thermally conductivepin fin.
 6. The apparatus according to claim 1 wherein said toric bodyis made of a thermally insulating material.
 7. The apparatus accordingto claim 1, including an outer surface housing said toric body.
 8. Theapparatus according to claim 7, wherein said outer surface supports saidheat sink above said toric body.
 9. The apparatus according to claim 8,wherein said outer surface is connected to both said heat spreader plateand said heat sink.
 10. The apparatus according to claim 1, including aheat source in thermal contact with said thermal spreader plate.
 11. Theapparatus according to claim 10, wherein said heat source is anelectronic device.